Demolition hammers and other such impact devices have been utilized for many years to deliver impact blows to an impact tool for the purpose of breaking up paving materials and other related activities requiring impact blows. Prior art impact devices utilizing high powered air compressors can be described as using the brute force approach to delivering impact blows. The requirement for a large compressor limits their mobility and access to many areas. In addition, the noise from the large air compressor and air exhausted from the air driven hammer is excessive and the fumes from the large internal combustion engine add to air polution.
Rotating shaft driven hammers such as electric motor hammers have failed to overcome all of the deficiencies related to the use of air compressor devices. Electric motor hammers of the prior art have had a relatively short service life and are not capable of delivering high power impact blows. In such prior art devices, the structure must be sufficiently light so that the device may be handled and manipulated manually. Yet, the presently devised light structures cannot carry the higher loads necessary to convert this power to performance. Accordingly, the current practice is to restrict the performance to a small fraction of what could otherwise be obtained so that the device will have a reasonably long service life. Further, the power efficiency of currently available devices is extremely low. In a typical impact device, only 10% of the power input to the device is actually utilized in breaking concrete.
Prior art devices utilizing a rotary shaft drive for the exciter means may be represented as being one of two types.
In the first type, the drive motor is a relatively large portion of the mass of the frame and its attachments and is mounted offside from the line of action of the main reaction force which drives the ram or piston to accomplish the impacting. In such arrangements the center of mass of the frame and attachments secured thereto are not aligned with the reactive force from the impacting piston. During operation, the high magnitude of these oscillating reactive forces induces extraneous oscillating torques that cause extraneous rotations of the frame as a whole as well as extraneous deflections and vibrations in the structure itself. These extraneous motions dissipate considerable power and shorten service life unnecessarily.
In the second arrangement common in prior art devices, the motor is not offset, but the mechanism including gears and their shafts, which carry the high reactive oscillating forces, must be supported by the supporting frame. The supporting frame is relatively light of necessity and hence is resilient to these oscillating forces. This arrangement also induces extraneous and undesirable forces and torques in the crank arm, shaft, and its support, which are for all prior art devices, unsymmetrically located relative to the line of action of the main reactive force from the impacting piston.
With both common prior art arrangements then, the high reactive oscillating force which is oscillating or impacting against the resilient supporting frame causes it to oscillate and deflect in numerous complex and high magnitude motions. Furthermore, such oscillatory deflections in the supporting frame will "reflect" back onto the driving mechanism in such a way as to complicate the motions and forces further.
In prior art devices, the standard analytical procedures failed to take adequate account of the extraneous or secondary effects. With this approach, the value of the oscillating force can be represented by EQU F.sub.r = A.sub.o sin 2.pi. ft 1.
where A.sub.o is the maximum amplitude, t the time in seconds, and f the frequency of the oscillatory force in cycles per second.
In fact, however, because of the presence of the extraneous effects as described above, the actual oscillating force is not this simple sinusoidal relation, but rather is more correctly represented by a series of functions such as EQU F.sub.r = A.sub.o sin 2 .pi. ft + A.sub.1 sin 4.pi. ft + A.sub.2 sin 6.pi. ft + . . . 2.
This equation is what is referred to in mathematics as a Fourier series and constitutes a frequency analysis of the actual wave form of the reactive force. A typical experimental curve is shown in FIG. 9. This curve shows the time variation of the force at one point of a typical structure under the action of a typical oscillating force. The curve for the structure motion or deflections is generally similar in shape to that of the force itself.
According to Equation 2, in addition to the fundamental component, that is the first term only, which is represented as in Equation 1, there are present many other oscillation frequencies. Each of these has an amplitude (A.sub.i) or magnitude that is induced by and determined by the factors discussed above.
Whereas an exact representation of the curve for any particular device will be dependent upon the mass distribution of the components and frame as well as their size, shape and elasticity of the material of which they are constructed, it is sufficient to note that the larger the disturbing causes such as offsets and asymmetries of components and their supports, and the necessity to support these in the frame as discussed above, the larger the amplitudes (A.sub.i) of these components in both force and deflections.
While these additional forces are referred to as extraneous and secondary, they are actually of utmost importance. The importance of these extraneous effects becomes apparent when it is observed relatively how much power is dissipated by them. These forces include all the vibrations except those represented by the first term of Equation 2 and they do no work but shake the frame and its components as well as the operator. The power dissipated by a vibration is proportional not only to its amplitude and its dissipation constant, but also to the cube of its frequency. Therefore, it becomes apparent that these higher frequency extraneous vibrations can be a source of much power loss and much damage to the structure.
For example, in one instance experimentally observed by an oscilloscope, one component, the sixth harmonic, stood out. It had an amplitude of only 10% of the fundamental, but its frequency was six times as high. Accordingly, if the ratio of dissipation factor on this harmonic is only 10% of the dissipation factor on the fundamental, the power lost through it by dissipation from this harmonic is 0.1 .times. 0.1 .times. 6 .times. 6 .times. 6, or 2.16. In other words, there is over twice as much power dissipated through this "extraneous" vibration as is dissipated in the device accomplishing the work it was intended to do.
This analysis represents only one component and frequently numerous such components exist with frequencies from twice, to many times that of the fundamental. These extraneous forces build up such large power losses and tend to heat up the bearings causing fatigue in the linkage and supports as well as health damage to the operator.
The problem is compounded in prior art devices in that the undesirable disturbances are increased through the relatively short length connecting rod utilized. In these devices the linear acceleration, a, of the cylinder in its reciprocal action if the crank arm were rotating uniformly would be given by the expression EQU a = w.sup.2 r.sub.c cos wt + secondary terms 3.
where w is the angular frequency in radians per second and r.sub.c is the crank arm radius. The first term in this equation would produce a reaction force such as is given by Equation 1 discussed hereinbefore. The secondary terms decrease when the ratio L/r.sub.c of connecting rod length to crank arm radius is large. The effect of these secondary terms is to add harmonic components to the fundamental given by the Equation 1 as discussed before in connection with Equation 2 and to introduce lateral forces of an amplitude proportional to the ratio r.sub.c /L times the reaction force. Both of these effects are insignificant in the ratio L/r.sub.c is greater than 6. In substantially all prior art devices this ratio is less than 6.
Most prior art devices utilize a piston cylinder to cushion the impact of the tool bit in the mechanism. Frequently two circumferential rings of vents on the cylinder are utilized with the vents located with one ring located a distance equal to approximately one third of the cylinder length from one end of the cylinder and a second ring of vents equally spaced from the other end of the cylinder. In particular, the two rings are spaced apart a distance greater than the length of the piston in the axial direction where the piston bears against the cylinder.
A second approach utilized in prior art devices incorporates a series of vents or slots serving the same purpose and extending approximately along the middle one third of the cylinder length. The length of these series of vents or slots is also somewhat greater than the length of the piston. The typical function of these slots or vents is the act as a bypass between the trapped air at opposite ends of the cylinders so that the pressure therein is maintained at approximately the same value. Only after the piston closes off the vents entirely at the extremes of its travel is the air entrapped and compressed to serve as a cushion so that the piston will not impact by direct metal contact with either cylinder end. To accomplish this purpose, considerable open area for the vents and considerable air flow is required. Much power is dissipated in producing such air flow and when the piston is passing between the rings of the vents or extremes of the slots, the piston is slowed down, since the energy needed for the pumping and for friction is taken from the pistons' kinetic energy. Accordingly the velocity at impact of the piston shaft on the bit is lowered, reducing the effectiveness of the impact.
As the piston closes the vent at one end, a mass of air with an initial volume of about one third of that of the entire cylinder, is entrapped in that end and forms a cushion to soften the blow between the piston and cylinder end as the piston is decelerated and given a reverse velocity. Since the velocity reversal and downward acceleration necessary for inducing the high impact by the foregoing arrangement must be carried out in somewhat less than one third of the cylinder length, the piston can only compress the entrapped gas which constitutes the cushion in about one half to two thirds of this remaining length, or less than about one sixth or two ninths of the total cylinder length. Accordingly, there will be a force induced on the piston which varies with its travel relative to the cylinder length. Curve A in FIG. 10 is typically representative of this force relationship. The piston travel is shown in terms of a cylinder length h. This is the total average open height of the cylinder on one side of the piston. h equals the open volume V on that side of the piston divided by the piston area A. x is the actual piston excursion. x/h is 0 when the piston excursion is 0 at the point of zero force and 1.0 if the piston were to move so that the remaining volume on that side toward which it is displaced were 0. h is usually the same on each side of a double acting air spring.
A third prior configuration makes use of a long piston. With this configuration, as little as 5% of the cylinder initial volume is available for deceleration of the piston at the end of each stroke. Such high deceleration results in high cylinder pressures, as high as 20-25 atmospheres, and results in a very sharp peak in the force curve; therefore transferring high forces to the mechanism and frame and using a high ratio of piston excursion to total open cylinder length. Such ratio expressed, as a function of piston excursion x to cylinder length h is illustrated as curve B in FIG. 10.
Prior art devices therefore must operate on the curves such as A and B and therefore reach x/h ratios in excess of 0.7 to obtain the high return force necessary to decelerate the piston in the minimal distance provided. Operation in this range produces high harmonics such as are identified in Equation 2 and therefore places severe loads on the bearings and structure, passing along sharp reaction forces to the operator holding the equipment, and causing severe heating and heat losses. The characteristics of prior art devices are understandable in that the piston-cylinder is intended to provide an air cushion rather than to provide other functions, and is inherent in the basic gas law equation EQU FV.sup.n = AP.sub.i V.sub.i.sup.n 4.
where A is the piston area, P.sub.i is the initial pressure and V.sub.i is the initial volume of the piston, and n is a constant of value approximately 1.3. Assuming the normal thermo-dynamic relation would obtain where EQU T.sub.2 = T.sub.1 (P.sub.2 /P.sub. P.sub.1) .sup.(n-1)/n 5.
and T.sub.1 is the ambient temperature in degrees absolute, T.sub.2 is the final temperature, for a typical example with a compression ratio P.sub.2 /P.sub. P.sub.1 of 25 the temperature would rise to approximately 323.degree. C above ambient temperature. While this high temperature occurs only at the peak of compression, and the time average over the cycle of operation would be somewhat less, this average is still a very high value. The heat and consequent power loss by radiation and conduction will be considerable and hence the power efficiency of operation will be lowered accordingly, and the piston cylinder will heat up introducing difficult lubrication and materials problems.
Due to the aforedescribed deficiencies of prior art devices, an improved impact device is much sought after. Thus it would be advantageous to have an impact device with greater portability than prior art devices, and one having the capability of delivering heavy impact blows more efficiently and effectively with much reduced noise levels and with reduced air polution. Additionally, it would be advantageous to have a device which delivered less harmful oscillations or vibration to the operator and one which is less susceptible to breakdown in service.